US20130167957A1 - Hydraulic piston pump with a variable displacement throttle mechanism - Google Patents
Hydraulic piston pump with a variable displacement throttle mechanism Download PDFInfo
- Publication number
- US20130167957A1 US20130167957A1 US13/343,436 US201213343436A US2013167957A1 US 20130167957 A1 US20130167957 A1 US 20130167957A1 US 201213343436 A US201213343436 A US 201213343436A US 2013167957 A1 US2013167957 A1 US 2013167957A1
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- United States
- Prior art keywords
- pump
- cylinders
- inlet
- recited
- variable orifice
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
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Classifications
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
- F04B1/04—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement
- F04B1/0404—Details or component parts
- F04B1/0408—Pistons
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
- F04B1/04—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement
- F04B1/0404—Details or component parts
- F04B1/0421—Cylinders
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B1/00—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders
- F04B1/04—Multi-cylinder machines or pumps characterised by number or arrangement of cylinders having cylinders in star- or fan-arrangement
- F04B1/0404—Details or component parts
- F04B1/0426—Arrangements for pressing the pistons against the actuated cam; Arrangements for connecting the pistons to the actuated cam
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B49/00—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00
- F04B49/22—Control, e.g. of pump delivery, or pump pressure of, or safety measures for, machines, pumps, or pumping installations, not otherwise provided for, or of interest apart from, groups F04B1/00 - F04B47/00 by means of valves
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
- F04B53/00—Component parts, details or accessories not provided for in, or of interest apart from, groups F04B1/00 - F04B23/00 or F04B39/00 - F04B47/00
- F04B53/16—Casings; Cylinders; Cylinder liners or heads; Fluid connections
-
- Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
- Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
- Y10T—TECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
- Y10T137/00—Fluid handling
- Y10T137/8593—Systems
- Y10T137/85978—With pump
- Y10T137/85986—Pumped fluid control
Definitions
- the present invention relates to hydraulic pumps, such as those that have pistons that move radially against an eccentric shaft, and more particularly to mechanisms for controlling the flow of fluid through the cylinders in which the pistons move.
- a common type of radial piston pump comprises a body with a plurality of cylinders radially disposed around a drive shaft that is rotated by an external motor or engine.
- a separate piston is slideably received within each cylinder, thereby defining a chamber at the interior of the cylinder.
- the drive shaft has an eccentric cam and the pistons are biased by springs to ride against that cam.
- the pistons slide reciprocally within the respective cylinders, thereby reducing and expanding the volume of the cylinder chambers in a cyclical manner. The smallest volume occurs at the top dead center point of the piston cycle and the largest volume occurs at the bottom dead center point.
- An inlet port supplies fluid to an inlet passage that has a separate inlet into each cylinder. Every cylinder also has an outlet that is coupled by a separate outlet check valve to an outlet passage that leads to the outlet port of the pump.
- U.S. Pat. No. 3,434,428 discloses a pump of this configuration. The pump in that patent also has a throttle plate with apertures associated with the inlets for the cylinders. The throttle plate is rotated by an actuator to vary alignment of the apertures with the inlets and thereby alter the amount of fluid flowing between the common inlet passage and each cylinder inlet.
- a drawback of this type of pump is that during a dead portion of the piston cycle, between bottom dead center point and when the inlet becomes closed, no pumping action occurs. Specifically, fluid is neither being expelled from the cylinder nor being drawn into the cylinder during that dead portion, which can be a third of the piston cycle as shown in FIG. 6 of the U.S. Pat. No. 3,434,428. This inactive time and initial short sealing length results in a sizeable inefficiency. In addition this type of pump requires a relatively long piston stroke to accommodate the dead portion of the piston cycle, which increases the diameter of the pump.
- a pump includes a cylinder block with an inlet port, an outlet port, a plurality of cylinders disposed radially in the cylinder block.
- a plurality of inlet passages are each connected between the inlet port and a different one of the plurality of cylinders, and a plurality of outlet passages each connected between the outlet port and a different one of the plurality of cylinders.
- a separate piston is slideably located in a each of the plurality of cylinders and drive shaft is rotatably received in the cylinder block for driving the piston reciprocally the cylinders.
- a separate inlet check valve is located in each of the plurality of inlet passages and allows fluid flow only in a direction from the inlet port into one of the plurality of cylinders.
- a separate outlet check valve is located in each of the plurality of outlet passages and allow fluid flow only in a direction from one of the plurality of cylinders into the outlet port.
- a throttle plate communicates with each of the plurality of inlet passages for varying a rate of fluid flow through the inlet passages.
- the throttle plate extends across each of the plurality of inlet passages and has a plurality of control apertures there through.
- the throttle plate is moveable to alter alignment of the control apertures with the inlet passages and thereby vary a cross sectional area through which fluid flows in the inlet passages. This provides a variable orifice in each inlet passage.
- the flow area of the variable orifice is directly related to the magnitude of the fluid flow there through.
- the average rate of change of the flow area of the variable orifice relative to movement of the throttle plate is greater during a first half of the travel distance between the fully open and fully closed positions.
- the flow area of the variable orifice decreases at least 80 percent in the first half of the throttle plate travel distance from the fully open position. This rapid closure rate of the variable orifice occurs in what is referred to as the first section of the throttle plate rotation. Thereafter, the rate of change of the flow area decreases significantly slower, requiring that the throttle plate move through the second half of travel distance to reduce the flow area the remaining 20 percent.
- FIG. 1 is a radial cross section showing the arrangement of the cylinders and pistons in the pump
- FIG. 2 is an axial cross section through the radial piston pump along line 2 - 2 in FIG. 2 ;
- FIG. 3 is an radial cross section through the radial piston pump along line 3 - 3 in FIG. 2 showing a position of a throttle plate in which apertures therein are in fully open states;
- FIG. 4 illustrates another position of a throttle plate in which the apertures are in partially open states
- FIG. 5 shows a further position of a throttle plate in which the apertures are closed
- FIG. 6 is an radial cross section through the radial piston pump similar to FIG. 5 , but showing an alternative arrangement of the apertures in the throttle plate;
- FIG. 7 is schematic diagram of a hydraulic circuit for controlling the position of the throttle plate.
- FIG. 8 is a graph of the relationship of the size of the open area of the apertures versus the position of the throttle plate.
- a hydraulic pump 10 has a cylinder block 30 with exterior first and second end surfaces 21 and 22 between which a cylindrical exterior side surface 38 extends.
- the cylinder block 30 has an inlet port 28 and an outlet port 29 through which hydraulic fluid is received and expelled from a hydraulic system.
- the inlet and outlet ports 28 and 29 open into inlet and outlet galleries 31 and 32 , respectively, that extend in circles through the cylinder block around a central shaft bore 41 in the cylinder block 30 .
- Three cylinders 36 extend radially outward from and are oriented at 120 degree increments around the central shaft bore 41 .
- Each cylinder 36 comprises a tubular sleeve 39 that is inserted into a bore in the cylinder block 30 .
- the sleeve 39 is beneficial in reducing the diameter of the pump 10 as will be described, the sleeve can be eliminated by using a material for the cylinder block that can be machined to form the cylinder bores.
- Each cylinder 36 has an opening through the cylindrical side surface 38 of the cylinder block 30 .
- a sealing cup 24 with an O-ring is placed inside each opening and a continuous band-shaped closing ring 35 extends around the side surface 38 tightly closing each of the cylinder openings.
- the closing ring 35 eliminates the relatively long plugs that projected outward from the cylinders in conventional pump designs and thereby reduces the overall diameter of the pump 10 .
- a plurality of inlet passages 26 are formed by first bores that extend into the first end surface 21 of the cylinder block 30 and each inlet passage opens into both the inlet gallery 31 and a respective one of the cylinders 36 .
- each inlet passage 26 is directly connected to both the inlet gallery 31 and one of the cylinders 36 .
- a separate inlet check valve 33 is located in each of those inlet passages 26 . The inlet check valve 33 opens when the pressure within the inlet passage 26 is greater than the pressure within the associated cylinder chamber 37 , as occurs during the intake phase of the pumping cycle.
- a plurality of outlet passages 27 are formed by second bores that extend into the second end surface 22 of the cylinder block 30 with each outlet passage opening into both the outlet gallery 32 and a respective one of the cylinders 36 . Every outlet passage 27 is directly connected to both the outlet gallery 32 and one of the cylinders 36 .
- a separate outlet check valve 34 is located in each of those outlet passages 27 .
- the outlet check valve 34 opens when pressure within the associated cylinder chamber 37 is greater than the pressure within the outlet gallery 32 , as occurs during the exhaust phase of the pumping cycle. It should be understood that the inlet and outlet galleries 31 and 32 communicate with all the piston cylinders in the pump and an identical pair of check valves is provided for each cylinder.
- Each of the inlet and outlet check valves 33 and 34 is passive, meaning that it operates in response to pressure exerted thereon and not by an actuator, such as an electric solenoid.
- the tubular sleeve 39 that partially forms the cylinder 36 enables the inlet and outlet check valves 33 and 34 to be placed closer to the longitudinal axis 25 of the drive shaft 40 .
- the inlet and outlet check valves 33 and 34 are within the closed curved perimeter defined by the exterior side surface 38 of the cylinder block 30 .
- the valves had to be outward from the top dead center position of the piston in order to receive the fluid forced out of the cylinder chamber 37 .
- the tubular sleeve 39 extends partially over the opening between the cylinder chamber 37 and the bores in which the check valves 33 and 34 are located, thereby extending the cylinder bore farther into the cylinder chamber 37 .
- a drive shaft 40 extends through the shaft bore 41 and is rotatable therein being supported by a pair of bearings 42 .
- the center section of the drive shaft 40 within the cylinder block 30 has an eccentric cam 44 .
- the cam 44 has a circular outer surface, the center of which is offset from longitudinal axis 25 of the drive shaft 40 .
- a cam bearing 46 has an inner race 47 that is pressed onto the outer circumferential surface of the cam 44 and an outer race 48 .
- a plurality of rollers 49 are located between the inner race 47 and an outer race 48 of the cam bearing. With the proper heat treatment and surface finishing, the surface of the cam 44 can serve as the inner bearing race.
- the cam bearing 46 improves the efficiency of the pump 10 over previous pumps that used a sliding journal bearing for this function.
- the rollers may be cylindrical, spherical, or other shapes.
- a separate piston assembly 51 is slideably received within each of the cylinders 36 .
- Every piston assembly 51 comprises a piston 52 and a piston rod 54 .
- the piston rod 54 extends between the piston 52 and the cam bearing 46 .
- the piston rod 54 has a curved shoe 56 which abuts the outer race 48 of the cam bearing 46 .
- the shoe 56 is wider than the shaft of the piston rod creating a flange portion.
- a pair of annular retaining rings 58 extend around the cam 44 engaging the flange portion of each piston rod shoe 56 , thereby holding the piston rods 54 against the cam bearing 46 , which is particularly beneficial during the intake stroke portion of a pumping cycle.
- the retaining rings 58 eliminate the need for a spring to bias the piston assembly 51 against the cam bearing 46 .
- the curved shoe 56 evenly distributes the piston load over a wide area of the cam bearing 46 .
- the outer race 48 of the cam bearing 46 remains relatively stationary.
- the outer race 48 rotates at a very slow rate in comparison to the speed of the drive shaft and the inner race 47 . Therefore, there is little relative motion between each piston shoe 56 and the cam bearing's outer race 48 .
- the piston 52 is cup-shaped having an interior cavity 53 which opens toward the drive shaft 40 .
- An end of the piston rod 54 is received within that interior cavity 53 and has a partially spherical head 60 that fits into a mating partially spherical depression 62 in the piston 52 .
- the head of the piston 52 may have an aperture 50 there through to convey hydraulic fluid from the cylinder chamber 37 to lubricate the interface between a spherical head 60 and the piston 52 .
- the piston rod 54 is held against the piston 52 by an open single bushing or a split bushing 55 and a snap ring 57 that rests in an interior groove in the piston's interior cavity 53 . As the piston rod 54 follows the eccentric motion of the cam 44 and the piston 52 in turn follows by sliding within the cylinder 36 .
- the bushing and snap ring arrangement allows the spherical head 60 of the piston rod to pivot with respect to the piston 52 when a rotational moment is imposed onto the piston rod 54 by rotation of the cam 44 . Because of that pivoting, the rotational moment is not transferred into the piston 52 , thereby minimizing the lateral force between the piston and the wall of the cylinder 36 .
- the drive shaft 40 includes an internal lubrication passage 64 extending from one end to the outer surface of the cam 44 .
- the lubrication passage 64 has a single opening in that outer surface at the center of the eccentric apex of the cam to feed fluid into the cam bearing 46 .
- the other end of the lubrication passage 64 opens into a chamber 66 at the end of the drive shaft 40 and that chamber receives relatively low pressure fluid through a feeder passage 68 from the inlet gallery 31 .
- centrifugal force expels fluid from the lubrication passage 64 into the cam bearing 46 .
- This action draws additional fluid into the lubrication passage 64 from the chamber 66 , thereby providing a pumping function for fluid that lubricates the cam bearing 46 .
- the cam bearing 46 has an inner race 47 , that inner race has apertures that convey the lubricating fluid to the rollers 49 .
- the outer race 48 also has through holes to lubricate the shoes 56 of the piston rods 54 , thereby providing splash lubrication and eliminating a need to have the central shaft bore 41 filled with fluid. Not having the crankcase filled with fluid reduces windage drag on the eccentric cam 44 and improves efficiency of the pump.
- Additional lubricating passages 59 are provided to convey fluid from the shaft bore 41 to the bearings 42 for the drive shaft 40 .
- the fluid used for lubrication exits the central shaft bore 41 through a standard drain port 69 from which the fluid is conveyed to a tank for the hydraulic system.
- Rotation of the eccentric cam 44 causes each piston 52 to move cyclically within the respective cylinder 36 , away from the sealing cup 24 during a fluid intake phase and then toward the sealing cup 24 during a fluid exhaust phase. Because of the radial arrangement of the cylinders 36 , at any point in time some pistons 52 are in the intake phase while other pistons are in the exhaust phase.
- the piston 52 illustrated in FIG. 2 is at the a top dead center position when the volume of its cylinder chamber 37 is the smallest, which occurs at a transition point from the exhaust phase to the intake phase during each piston cycle. From this point, the outlet check valve 34 closes and further rotation of the eccentric cam 44 moves the piston 52 into the intake phase.
- the volume of the cylinder chamber 37 increases, thereby initially decompressing the fluid remaining therein which tends to drive or put energy back into the drive shaft 40 . Thereafter, further increase in the cylinder volume produces a negative gauge pressure therein.
- the inlet check valve 33 is forced open by a positive atmospheric pressure applied from the inlet gallery 31 .
- the hydraulic pump 10 includes a throttle mechanism that varies the inlet opening area from the shared inlet gallery 31 into the inlet passage 26 and through the inlet check valve 33 for each cylinder 36 during the intake phase.
- That throttle mechanism comprises a circular throttle plate 90 and an abutting transition plate 91 that are sandwiched between two sections of the cylinder block 30 so as to extend across each of the plurality of inlet passages 26 .
- the throttle plate 90 and the transition plate 91 have central apertures 92 and 93 , respectively through which the drive shaft 40 extends.
- the transition plate 91 is held stationary within the cylinder block 30 and has a plurality of transmission apertures 94 , each fixedly aligned with one of the inlet passages 26 .
- the throttle plate 90 is rotatable around the drive shaft 40 and has a plurality of control apertures 95 proximate to the transmission apertures 94 in the transition plate 91 .
- the control apertures 95 of the throttle plate 90 and the transmission apertures 94 in the transition plate 91 are formed on nearly the same radius as that of the inlet passages 26 , thus assuring registration of those apertures with the inlet passages upon rotation of the throttle plate through a predefined arc.
- rotation of the throttle plate aligns and misaligns the control apertures 95 with the transmission apertures 94 , thereby creating variable orifices that control the fluid flow between the inlet gallery 31 and the cylinders 36 .
- the hydraulic pump 10 further includes an actuator 100 for rotating the throttle plate 90 within the cylinder block 30 .
- a tab 98 projects outward from the outer edge of the throttle plate 90 and into an actuator bore 102 in the cylinder block 30 .
- the actuator bore 102 has a control port 104 to which a hydraulic conduit from a control circuit connects.
- An actuator piston 108 is slideably received in the actuator bore 102 and engages the tab 98 of the throttle plate 90 .
- Pressurized fluid applied to the control port 104 drives the piston to the right in the actuator bore 102 (see FIG. 3 ), thereby causing the throttle plate to rotate into different positions such as those shown in FIGS. 4 and 5 .
- FIG. 7 depicts one type of a hydraulic circuit 140 that controls the displacement of the pump 10 by rotating the throttle plate 90 to maintain a desired pressure at the outlet port 29 of the pump.
- the pump outlet port 29 is connected to a conventional control valve 105 that controls the operation of a hydraulic actuator 106 , such as a motor or a piston/cylinder actuator of a machine with which the pump 10 is used.
- the hydraulic circuit 140 responds to a standard load senses pressure signal LS, received from the hydraulic actuator 106 , by maintaining the displacement of the pump 10 to produce a desired output pressure for operating the hydraulic actuator.
- Other hydraulic circuits can be used to operate the throttle plate actuator 100 .
- the angular position of the throttle plate 90 within the cylinder block 30 determines the alignment of the control apertures 95 in the throttle plate with the transmission apertures 94 in the transition plate 91 . Varying that alignment alters the degree to which those apertures overlap and thus alters the cross sectional area through which fluid is able to flow between the inlet gallery 31 and the cylinders 36 during the piston cycle intake phase.
- the adjustable alignment of the transmission and control apertures 94 and 95 forms a variable orifice in that flow path provided by the inlet passages 26 .
- Both the control apertures 95 and the transmission apertures 94 have unique shapes so that fluid flow varies in a specific manner to regulate the displacement of the pump 10 and maintain the output pressure at a desired level.
- FIG. 3 illustrates the control apertures 95 and the transmission apertures 94 in a fully aligned orientation that provides the maximum flow between the inlet gallery 31 and cylinders 36 .
- Transverse cross section as used herein means a cross section across a control aperture in a plane that is transverse to the direction that fluid flows through the aperture.
- each control aperture 95 has a transverse cross sectional shape that has an oval primary region 96 from which a tapered region 97 projects, like a beak of a bird, and terminates at an apex.
- the primary region 96 has a relatively large cross sectional area as compared to the cross sectional area of the tapered region 97 .
- the control apertures 95 can have other shapes and still attain variation of the rate of change of the fluid flow, as described herein.
- Each transmission aperture 94 in the transition plate 91 has a size and shape which ensures that the entire cross sectional area of the associated control aperture 95 communicates with the inlet passage 26 when the throttle plate 90 in the fully aligned position. That full alignment of the transmission and control apertures 94 and 95 enables the entire area of the control aperture 95 to conduct fluid through the throttle plate 90 and thus provides the maximum flow of fluid from the inlet gallery 31 into each cylinder 36 during the intake phase of the piston cycle.
- a spring 114 biases the actuator piston 108 into a position in which the throttle plate 90 is in the fully aligned aperture position.
- the amount of this flow can be proportionally controlled by controlling the rotational position of the throttle plate 90 and thus the amount of that aperture overlap.
- the tapered aperture regions 97 cause the flow area to change at a smaller rate than occurred during previous motion to reach that intermediate position from the fully aligned position of the transmission and control apertures 94 and 95 .
- an relatively smaller change in flow area occurs than happened previously. Therefore, the rate that the open area of the control apertures 95 changes decreases as that open area becomes smaller.
- FIG. 6 illustrates an second hydraulic pump 200 that is similar to the first hydraulic pump 10 as shown in FIG. 5 in which similar components have been assigned identical reference numerals.
- the distinction between the first and second hydraulic pumps is that the transmission apertures 202 in the transition plate 91 of the second hydraulic pump 200 have an oval primary section 206 from which a tapered section 208 projects, like a beak of a bird, and terminates at an apex.
- the control apertures 204 in the throttle plate 90 are identical to the transmission apertures 94 in the transition plate 91 of the first hydraulic pump 10 . In other words, the shapes of the transmission and the control apertures are switched in the second hydraulic pump 200 . Nevertheless, the transmission apertures 202 and the control apertures 204 function in the same manner, as described with respect to the first hydraulic pump 10 , regarding creating variable orifices that control the fluid flow between the inlet gallery and the cylinders of the pump.
- FIG. 8 graphically illustrates the relationship of the size of the open area, or flow area, of the control apertures 95 versus the position of the throttle plate 90 , which is related to the linear position of the actuator piston 108 for the exemplary pump.
- the flow area of the orifice is directly related to the magnitude of the fluid flow there through.
- the actuator piston and the throttle plate move from a first position, corresponding with the orifice being fully (100 percent) open, to a second position, corresponding with the orifice being fully closed (0 percent open).
- a mid position is located halfway between the first and second positions, i.e., at 50 percent of the travel distance between the first and second positions.
- the average rate of change of the flow area of the variable orifice relative to movement of the actuator piston is significantly greater during the first half of travel compared to the second half of travel.
- the flow area of the variable orifice created by the position of the control aperture 95 relative to the transmission aperture 94 decreases by at least 80 percent in the first 50 percent of the travel of the actuator piston from the initial first position, as depicted a point 122 that occurs at the mid position in FIG. 3 .
- This rapid closure rate of the variable orifice occurs in what is referred to as a first section 124 of the throttle plate rotation.
- the rate of change of the flow area decreases significantly slower, requiring that the actuator piston 108 move the remaining 50 percent of the travel distance (the second section 126 of the throttle plate rotation) to reduce the flow area the final 20 percent to the fully closed position.
- the piston and the throttle plate decreases the flow from 20 percent of the maximum flow to zero flow over the same amount (i.e., 50 percent) of throttle plate rotation as it takes to decrease the flow from 100 percent to 20 percent of the maximum flow.
- the flow area of the variable orifice changes from the maximum flow area to about 20 percent of that maximum flow area at a rate that is at least twice as fast and the rate at which the flow area changes between about 20 percent of that maximum flow area and zero flow area. Therefore from the fully aligned aperture position, rotation of the throttle plate initially produces a relatively rapid decrease in flow area and then the flow area decrease occurs at a slower rate the as aperture motion approached the closed position.
- the inverse rates of change occur as the throttle plate 90 moves clockwise in the drawings and the variable orifice, formed by the degree of alignment of the control apertures 95 with the transmission apertures 94 , opens greater amounts.
- a throttle plate 90 to control the amount of flow between the inlet gallery 31 and the inlet passages 26 enables the displacement of the pump 10 to be dynamically varied.
- the throttle plate apertures 95 are only partially aligned with the transition plate transmission apertures 94 , the amount of fluid flowing into the cylinder chamber 37 during the intake phase of each piston cycle is reduced.
- the piston 52 reaches bottom dead center without the cylinder chamber 37 being completely filled with hydraulic fluid.
- the amount of lost displacement does not vary significantly as a function of the speed of the pump, since the average pressure drop across the throttle plate is constant for typical pump speeds of 800 to 2500 RPM.
- the present pump configuration with the rotatable throttle plate 90 provides variable throttle choking at the input of each inlet check valve. This has a significant advantage over a pump that has throttle choking at a single place for all the cylinders, such as between the inlet port 28 and the inlet gallery 31 .
- the per inlet check valve choking arrangement of the present pump 10 the fluid volume between the throttle plate and the inlet check valve is relatively small and results in improved consistency and dynamic response in both starting and stopping fluid flow.
Abstract
Description
- Not Applicable
- Not Applicable
- 1. Field of the Invention
- The present invention relates to hydraulic pumps, such as those that have pistons that move radially against an eccentric shaft, and more particularly to mechanisms for controlling the flow of fluid through the cylinders in which the pistons move.
- 2. Description of the Related Art
- A common type of radial piston pump comprises a body with a plurality of cylinders radially disposed around a drive shaft that is rotated by an external motor or engine. A separate piston is slideably received within each cylinder, thereby defining a chamber at the interior of the cylinder. The drive shaft has an eccentric cam and the pistons are biased by springs to ride against that cam. As the cam rotates, the pistons slide reciprocally within the respective cylinders, thereby reducing and expanding the volume of the cylinder chambers in a cyclical manner. The smallest volume occurs at the top dead center point of the piston cycle and the largest volume occurs at the bottom dead center point.
- An inlet port supplies fluid to an inlet passage that has a separate inlet into each cylinder. Every cylinder also has an outlet that is coupled by a separate outlet check valve to an outlet passage that leads to the outlet port of the pump. U.S. Pat. No. 3,434,428 discloses a pump of this configuration. The pump in that patent also has a throttle plate with apertures associated with the inlets for the cylinders. The throttle plate is rotated by an actuator to vary alignment of the apertures with the inlets and thereby alter the amount of fluid flowing between the common inlet passage and each cylinder inlet.
- With this type of pump, as the piston moves from the top dead center point, fluid is not initially drawn into the expanding cylinder chamber because the location of the piston blocks the inlet. The piston has to move a considerable distance from the top dead center point before the inlet is unblocked and fluid from the inlet passage is drawn into the expanding cylinder chamber. After the bottom dead center point, the volume of the cylinder chamber begins reducing, however, the inlet still is open which prevents outlet check valve from opening. Here too, the piston must move some distance before the piston blocks the inlet and causes pressure in the cylinder chamber to increase. As the piston starts to pump, the sealing land of the piston is low in the cylinder and high pressure fluid leakage occurs thereby making this form of aspiration initially in-efficient. Eventually the pressure rises to a level that forces the outlet valve to open an outlet path through which the fluid is exhausted from the cylinder chamber. That exhausting continues until the piston again reaches the top dead center point.
- A drawback of this type of pump is that during a dead portion of the piston cycle, between bottom dead center point and when the inlet becomes closed, no pumping action occurs. Specifically, fluid is neither being expelled from the cylinder nor being drawn into the cylinder during that dead portion, which can be a third of the piston cycle as shown in FIG. 6 of the U.S. Pat. No. 3,434,428. This inactive time and initial short sealing length results in a sizeable inefficiency. In addition this type of pump requires a relatively long piston stroke to accommodate the dead portion of the piston cycle, which increases the diameter of the pump.
- These prior radial piston pumps also had a relatively large diameter due to the outlet valves and the outlet passage being located radially outward from each cylinder. For many machines, the amount of space for the pump is limited, thus it is desirable to reduce the size of the pump. More specifically, many times the pump is mounted alongside an engine or transmission and the radial space is limited preventing the installation of typical radial piston pumps.
- A pump includes a cylinder block with an inlet port, an outlet port, a plurality of cylinders disposed radially in the cylinder block. A plurality of inlet passages are each connected between the inlet port and a different one of the plurality of cylinders, and a plurality of outlet passages each connected between the outlet port and a different one of the plurality of cylinders. A separate piston is slideably located in a each of the plurality of cylinders and drive shaft is rotatably received in the cylinder block for driving the piston reciprocally the cylinders.
- A separate inlet check valve is located in each of the plurality of inlet passages and allows fluid flow only in a direction from the inlet port into one of the plurality of cylinders. A separate outlet check valve is located in each of the plurality of outlet passages and allow fluid flow only in a direction from one of the plurality of cylinders into the outlet port.
- A throttle plate communicates with each of the plurality of inlet passages for varying a rate of fluid flow through the inlet passages. In one embodiment, the throttle plate extends across each of the plurality of inlet passages and has a plurality of control apertures there through. The throttle plate is moveable to alter alignment of the control apertures with the inlet passages and thereby vary a cross sectional area through which fluid flows in the inlet passages. This provides a variable orifice in each inlet passage.
- One aspect of the present pump is that the flow area of the variable orifice is directly related to the magnitude of the fluid flow there through. Broadly speaking, as the throttle plate is moved from a position corresponding with the variable orifice being fully open to a position with the variable orifice is fully closed, the average rate of change of the flow area of the variable orifice relative to movement of the throttle plate is greater during a first half of the travel distance between the fully open and fully closed positions. For example, the flow area of the variable orifice decreases at least 80 percent in the first half of the throttle plate travel distance from the fully open position. This rapid closure rate of the variable orifice occurs in what is referred to as the first section of the throttle plate rotation. Thereafter, the rate of change of the flow area decreases significantly slower, requiring that the throttle plate move through the second half of travel distance to reduce the flow area the remaining 20 percent.
-
FIG. 1 is a radial cross section showing the arrangement of the cylinders and pistons in the pump; -
FIG. 2 is an axial cross section through the radial piston pump along line 2-2 inFIG. 2 ; -
FIG. 3 is an radial cross section through the radial piston pump along line 3-3 inFIG. 2 showing a position of a throttle plate in which apertures therein are in fully open states; -
FIG. 4 illustrates another position of a throttle plate in which the apertures are in partially open states; -
FIG. 5 shows a further position of a throttle plate in which the apertures are closed; -
FIG. 6 is an radial cross section through the radial piston pump similar toFIG. 5 , but showing an alternative arrangement of the apertures in the throttle plate; -
FIG. 7 is schematic diagram of a hydraulic circuit for controlling the position of the throttle plate; and -
FIG. 8 is a graph of the relationship of the size of the open area of the apertures versus the position of the throttle plate. - The term “directly connected” as used herein means that the associated components are connected together by a conduit without any intervening element, such as a valve, an orifice or other device, which restricts or controls the flow of fluid beyond the inherent restriction of any conduit. References herein to directional relationships and movement, such as top and bottom or left and right, refer to the relationship and movement of the components in the orientation illustrated in the drawings, which may not be the orientation of the components as attached to machinery.
- With reference to
FIGS. 1 and 2 , ahydraulic pump 10 has acylinder block 30 with exterior first andsecond end surfaces exterior side surface 38 extends. Thecylinder block 30 has aninlet port 28 and anoutlet port 29 through which hydraulic fluid is received and expelled from a hydraulic system. The inlet andoutlet ports outlet galleries cylinder block 30. Threecylinders 36 extend radially outward from and are oriented at 120 degree increments around the central shaft bore 41. Although theexemplary pump 10 is illustrated with three cylinders to simplify the drawings, in practice the pump may have a greater number of cylinders (e.g., 6 or 8 cylinders) to reduce torque, flow and pressure ripples at the outlet. Eachcylinder 36 comprises atubular sleeve 39 that is inserted into a bore in thecylinder block 30. Although thesleeve 39 is beneficial in reducing the diameter of thepump 10 as will be described, the sleeve can be eliminated by using a material for the cylinder block that can be machined to form the cylinder bores. Eachcylinder 36 has an opening through thecylindrical side surface 38 of thecylinder block 30. A sealingcup 24 with an O-ring is placed inside each opening and a continuous band-shapedclosing ring 35 extends around theside surface 38 tightly closing each of the cylinder openings. Theclosing ring 35 eliminates the relatively long plugs that projected outward from the cylinders in conventional pump designs and thereby reduces the overall diameter of thepump 10. - With particular reference to
FIG. 2 , a plurality ofinlet passages 26 are formed by first bores that extend into thefirst end surface 21 of thecylinder block 30 and each inlet passage opens into both theinlet gallery 31 and a respective one of thecylinders 36. In other words, eachinlet passage 26 is directly connected to both theinlet gallery 31 and one of thecylinders 36. A separateinlet check valve 33 is located in each of thoseinlet passages 26. Theinlet check valve 33 opens when the pressure within theinlet passage 26 is greater than the pressure within the associatedcylinder chamber 37, as occurs during the intake phase of the pumping cycle. A plurality ofoutlet passages 27 are formed by second bores that extend into thesecond end surface 22 of thecylinder block 30 with each outlet passage opening into both theoutlet gallery 32 and a respective one of thecylinders 36. Everyoutlet passage 27 is directly connected to both theoutlet gallery 32 and one of thecylinders 36. A separateoutlet check valve 34 is located in each of thoseoutlet passages 27. Theoutlet check valve 34 opens when pressure within the associatedcylinder chamber 37 is greater than the pressure within theoutlet gallery 32, as occurs during the exhaust phase of the pumping cycle. It should be understood that the inlet andoutlet galleries outlet check valves - The
tubular sleeve 39 that partially forms thecylinder 36 enables the inlet andoutlet check valves longitudinal axis 25 of thedrive shaft 40. Note that the inlet andoutlet check valves exterior side surface 38 of thecylinder block 30. In prior configurations the valves had to be outward from the top dead center position of the piston in order to receive the fluid forced out of thecylinder chamber 37. As shown inFIG. 2 , thetubular sleeve 39 extends partially over the opening between thecylinder chamber 37 and the bores in which thecheck valves cylinder chamber 37. - Referring again to both to
FIGS. 1 and 2 , adrive shaft 40 extends through the shaft bore 41 and is rotatable therein being supported by a pair ofbearings 42. The center section of thedrive shaft 40 within thecylinder block 30 has aneccentric cam 44. Thecam 44 has a circular outer surface, the center of which is offset fromlongitudinal axis 25 of thedrive shaft 40. As a consequence, as thedrive shaft 40 rotates within thecylinder block 30, thecam 44 rotates in an eccentric manner about theaxis 25 of the drive shaft. As specifically shown inFIG. 2 , acam bearing 46 has aninner race 47 that is pressed onto the outer circumferential surface of thecam 44 and anouter race 48. A plurality ofrollers 49 are located between theinner race 47 and anouter race 48 of the cam bearing. With the proper heat treatment and surface finishing, the surface of thecam 44 can serve as the inner bearing race. Thecam bearing 46 improves the efficiency of thepump 10 over previous pumps that used a sliding journal bearing for this function. The rollers may be cylindrical, spherical, or other shapes. - A
separate piston assembly 51 is slideably received within each of thecylinders 36. Everypiston assembly 51 comprises apiston 52 and apiston rod 54. Thepiston rod 54 extends between thepiston 52 and thecam bearing 46. Thepiston rod 54 has acurved shoe 56 which abuts theouter race 48 of thecam bearing 46. Theshoe 56 is wider than the shaft of the piston rod creating a flange portion. A pair of annular retaining rings 58 extend around thecam 44 engaging the flange portion of eachpiston rod shoe 56, thereby holding thepiston rods 54 against the cam bearing 46, which is particularly beneficial during the intake stroke portion of a pumping cycle. The retaining rings 58 eliminate the need for a spring to bias thepiston assembly 51 against thecam bearing 46. Thecurved shoe 56 evenly distributes the piston load over a wide area of thecam bearing 46. As thedrive shaft 40 andcam 44 rotate within thecylinder block 30, theouter race 48 of the cam bearing 46 remains relatively stationary. Theouter race 48 rotates at a very slow rate in comparison to the speed of the drive shaft and theinner race 47. Therefore, there is little relative motion between eachpiston shoe 56 and the cam bearing'souter race 48. - The
piston 52 is cup-shaped having aninterior cavity 53 which opens toward thedrive shaft 40. An end of thepiston rod 54 is received within thatinterior cavity 53 and has a partiallyspherical head 60 that fits into a mating partiallyspherical depression 62 in thepiston 52. The head of thepiston 52 may have anaperture 50 there through to convey hydraulic fluid from thecylinder chamber 37 to lubricate the interface between aspherical head 60 and thepiston 52. Thepiston rod 54 is held against thepiston 52 by an open single bushing or asplit bushing 55 and asnap ring 57 that rests in an interior groove in the piston'sinterior cavity 53. As thepiston rod 54 follows the eccentric motion of thecam 44 and thepiston 52 in turn follows by sliding within thecylinder 36. The bushing and snap ring arrangement allows thespherical head 60 of the piston rod to pivot with respect to thepiston 52 when a rotational moment is imposed onto thepiston rod 54 by rotation of thecam 44. Because of that pivoting, the rotational moment is not transferred into thepiston 52, thereby minimizing the lateral force between the piston and the wall of thecylinder 36. - With continuing reference to
FIG. 2 , thedrive shaft 40 includes aninternal lubrication passage 64 extending from one end to the outer surface of thecam 44. Thelubrication passage 64 has a single opening in that outer surface at the center of the eccentric apex of the cam to feed fluid into thecam bearing 46. The other end of thelubrication passage 64 opens into achamber 66 at the end of thedrive shaft 40 and that chamber receives relatively low pressure fluid through afeeder passage 68 from theinlet gallery 31. As thedrive shaft 40 rotates, centrifugal force expels fluid from thelubrication passage 64 into thecam bearing 46. This action draws additional fluid into thelubrication passage 64 from thechamber 66, thereby providing a pumping function for fluid that lubricates thecam bearing 46. If the cam bearing 46 has aninner race 47, that inner race has apertures that convey the lubricating fluid to therollers 49. Theouter race 48 also has through holes to lubricate theshoes 56 of thepiston rods 54, thereby providing splash lubrication and eliminating a need to have the central shaft bore 41 filled with fluid. Not having the crankcase filled with fluid reduces windage drag on theeccentric cam 44 and improves efficiency of the pump.Additional lubricating passages 59 are provided to convey fluid from the shaft bore 41 to thebearings 42 for thedrive shaft 40. The fluid used for lubrication exits the central shaft bore 41 through astandard drain port 69 from which the fluid is conveyed to a tank for the hydraulic system. - Rotation of the
eccentric cam 44 causes eachpiston 52 to move cyclically within therespective cylinder 36, away from the sealingcup 24 during a fluid intake phase and then toward the sealingcup 24 during a fluid exhaust phase. Because of the radial arrangement of thecylinders 36, at any point in time somepistons 52 are in the intake phase while other pistons are in the exhaust phase. - The
piston 52 illustrated inFIG. 2 is at the a top dead center position when the volume of itscylinder chamber 37 is the smallest, which occurs at a transition point from the exhaust phase to the intake phase during each piston cycle. From this point, theoutlet check valve 34 closes and further rotation of theeccentric cam 44 moves thepiston 52 into the intake phase. During the intake phase, the volume of thecylinder chamber 37 increases, thereby initially decompressing the fluid remaining therein which tends to drive or put energy back into thedrive shaft 40. Thereafter, further increase in the cylinder volume produces a negative gauge pressure therein. As a result, theinlet check valve 33 is forced open by a positive atmospheric pressure applied from theinlet gallery 31. Thus, fluid flows from theinlet gallery 31 through theinlet passage 26 and theinlet check valve 33 into the expandingcylinder chamber 37. At this time, when a negative pressure in thecylinder chamber 37, the pressure in theoutlet gallery 32 is positive due to either the flow output of the other cylinder chambers passing through a restriction or a static or dynamic load on the output. That pressure differential forces theoutlet check valve 34 closed against its valve seat. The intake phase continues until theeccentric cam 44 moves thatpiston 52 to the bottom dead center position, at which the volume ofcylinder chamber 37 is the greatest. Thus the bottom dead center position occurs at a transition in the piston cycle from the intake phase to the exhaust phase. - Thereafter, further rotation of the
eccentric cam 44 moves thepiston 52 into the exhaust phase during which the piston moves outward, away from thecenter axis 25. That motion initially compresses the fluid in thecylinder chamber 37, thereby increasing the pressure of that fluid. Soon the pressure in thecylinder chamber 37 is approximately that same as the pressure in theinlet passage 26, at which point the associated spring closes the inletfirst check valve 33. Eventually, the cylinder chamber pressure exceeds the pressure in theoutlet gallery 32 and forces theoutlet check valve 34 open, releasing the fluid from thecylinder chamber 37 into the outlet gallery and to theoutlet port 29. - When continued rotation of the
eccentric cam 44 moves thepiston 52 to the top dead center position shown inFIG. 2 , the exhaust phase is complete and thereafter the piston transitions into the intake phase of another pumping cycle. - Because the inlet and
outlet check valves - With reference to
FIGS. 2 and 3 , thehydraulic pump 10 includes a throttle mechanism that varies the inlet opening area from the sharedinlet gallery 31 into theinlet passage 26 and through theinlet check valve 33 for eachcylinder 36 during the intake phase. That throttle mechanism comprises acircular throttle plate 90 and anabutting transition plate 91 that are sandwiched between two sections of thecylinder block 30 so as to extend across each of the plurality ofinlet passages 26. Thethrottle plate 90 and thetransition plate 91 havecentral apertures drive shaft 40 extends. Thetransition plate 91 is held stationary within thecylinder block 30 and has a plurality oftransmission apertures 94, each fixedly aligned with one of theinlet passages 26. Thethrottle plate 90 is rotatable around thedrive shaft 40 and has a plurality ofcontrol apertures 95 proximate to thetransmission apertures 94 in thetransition plate 91. Thecontrol apertures 95 of thethrottle plate 90 and thetransmission apertures 94 in thetransition plate 91 are formed on nearly the same radius as that of theinlet passages 26, thus assuring registration of those apertures with the inlet passages upon rotation of the throttle plate through a predefined arc. As will be described, rotation of the throttle plate aligns and misaligns thecontrol apertures 95 with thetransmission apertures 94, thereby creating variable orifices that control the fluid flow between theinlet gallery 31 and thecylinders 36. - The
hydraulic pump 10 further includes anactuator 100 for rotating thethrottle plate 90 within thecylinder block 30. For that purpose, atab 98 projects outward from the outer edge of thethrottle plate 90 and into anactuator bore 102 in thecylinder block 30. The actuator bore 102 has acontrol port 104 to which a hydraulic conduit from a control circuit connects. Anactuator piston 108 is slideably received in the actuator bore 102 and engages thetab 98 of thethrottle plate 90. Pressurized fluid applied to thecontrol port 104 drives the piston to the right in the actuator bore 102 (seeFIG. 3 ), thereby causing the throttle plate to rotate into different positions such as those shown inFIGS. 4 and 5 . -
FIG. 7 depicts one type of ahydraulic circuit 140 that controls the displacement of thepump 10 by rotating thethrottle plate 90 to maintain a desired pressure at theoutlet port 29 of the pump. Thepump outlet port 29 is connected to aconventional control valve 105 that controls the operation of ahydraulic actuator 106, such as a motor or a piston/cylinder actuator of a machine with which thepump 10 is used. Thehydraulic circuit 140 responds to a standard load senses pressure signal LS, received from thehydraulic actuator 106, by maintaining the displacement of thepump 10 to produce a desired output pressure for operating the hydraulic actuator. Other hydraulic circuits can be used to operate thethrottle plate actuator 100. - The angular position of the
throttle plate 90 within thecylinder block 30 determines the alignment of thecontrol apertures 95 in the throttle plate with thetransmission apertures 94 in thetransition plate 91. Varying that alignment alters the degree to which those apertures overlap and thus alters the cross sectional area through which fluid is able to flow between theinlet gallery 31 and thecylinders 36 during the piston cycle intake phase. In other words, the adjustable alignment of the transmission andcontrol apertures inlet passages 26. Both thecontrol apertures 95 and thetransmission apertures 94 have unique shapes so that fluid flow varies in a specific manner to regulate the displacement of thepump 10 and maintain the output pressure at a desired level.FIG. 3 illustrates thecontrol apertures 95 and thetransmission apertures 94 in a fully aligned orientation that provides the maximum flow between theinlet gallery 31 andcylinders 36. As thethrottle plate 90 rotates counter clockwise and the transmission andcontrol apertures FIG. 4 . As the orifice area thereafter becomes smaller, the rate that the area changes decreases, i.e. the area changes more slowly for identical increments of change in the angular position of the throttle plate. - The variation in the rate of orifice area change is determined by the unique shape of the transverse cross section of the
control apertures 95 in thethrottle plate 90. Transverse cross section as used herein means a cross section across a control aperture in a plane that is transverse to the direction that fluid flows through the aperture. As shown inFIG. 3 , eachcontrol aperture 95 has a transverse cross sectional shape that has an ovalprimary region 96 from which a taperedregion 97 projects, like a beak of a bird, and terminates at an apex. Theprimary region 96 has a relatively large cross sectional area as compared to the cross sectional area of the taperedregion 97. Thecontrol apertures 95 can have other shapes and still attain variation of the rate of change of the fluid flow, as described herein. Eachtransmission aperture 94 in thetransition plate 91 has a size and shape which ensures that the entire cross sectional area of the associatedcontrol aperture 95 communicates with theinlet passage 26 when thethrottle plate 90 in the fully aligned position. That full alignment of the transmission andcontrol apertures control aperture 95 to conduct fluid through thethrottle plate 90 and thus provides the maximum flow of fluid from theinlet gallery 31 into eachcylinder 36 during the intake phase of the piston cycle. Aspring 114 biases theactuator piston 108 into a position in which thethrottle plate 90 is in the fully aligned aperture position. - From the fully aligned position in
FIG. 3 , application of pressurized fluid to thecontrol port 104 drives theactuator piston 108 which acts on thetab 98 rotating thethrottle plate 90 counter clockwise. Continued motion eventually moves thethrottle plate 90 into an intermediate position depicted inFIG. 4 . As thethrottle plate 90 moved between those positions the largerprimary regions 96 of thecontrol apertures 95 move over the edge of thetransmission apertures 94 in thetransition plate 91, thereby closing off some of the area of each control aperture. Because of the large size of the ovalprimary regions 96, the area through which fluid flows through the orifice, created by thecontrol apertures 95 and thetransmission apertures 94, diminishes at a relatively fast rate (seeFIG. 8 ). That is for a given incremental distance that theactuator piston 108 moves and thus for a given incremental angular change in throttle plate position, an relatively large change in flow occurs. - Upon reaching the intermediate position in
FIG. 4 , only the taperedregions 97 of thecontrol apertures 95 remain aligned to communicate with thetransmission apertures 94 in thetransition plate 91. Thus fluid can only flow through thethrottle plate 90 via those tapered sections. In this intermediate position, thecontrol apertures 95 are only partially aligned with thetransmission apertures 94 in thetransition plate 91. Depending upon the amount of overlap in this intermediate position, the amount of flow between theinlet gallery 31 and each of theinlet passages 26 is reduced from the fully aligned position. - The amount of this flow can be proportionally controlled by controlling the rotational position of the
throttle plate 90 and thus the amount of that aperture overlap. As the rotation of thethrottle plate 90 continues the taperedaperture regions 97 cause the flow area to change at a smaller rate than occurred during previous motion to reach that intermediate position from the fully aligned position of the transmission andcontrol apertures actuator piston 108 moves and for each given incremental angle change of the throttle plate, an relatively smaller change in flow area occurs than happened previously. Therefore, the rate that the open area of thecontrol apertures 95 changes decreases as that open area becomes smaller. - Continued activation of the
control actuator 100, results in thethrottle plate 90 eventually reaching the position illustrated inFIG. 5 in which thecontrol apertures 95 are entirely misaligned with thetransmission apertures 94 in thetransition plate 91. That is, no part of the throttleplate control apertures 95 overlaps or opens into the transitionplate transmission apertures 94 and fluid flow between theinlet gallery 31 and thecylinders 36 is blocked. -
FIG. 6 , illustrates an secondhydraulic pump 200 that is similar to the firsthydraulic pump 10 as shown inFIG. 5 in which similar components have been assigned identical reference numerals. The distinction between the first and second hydraulic pumps is that thetransmission apertures 202 in thetransition plate 91 of the secondhydraulic pump 200 have an ovalprimary section 206 from which atapered section 208 projects, like a beak of a bird, and terminates at an apex. Thecontrol apertures 204 in thethrottle plate 90 are identical to thetransmission apertures 94 in thetransition plate 91 of the firsthydraulic pump 10. In other words, the shapes of the transmission and the control apertures are switched in the secondhydraulic pump 200. Nevertheless, thetransmission apertures 202 and thecontrol apertures 204 function in the same manner, as described with respect to the firsthydraulic pump 10, regarding creating variable orifices that control the fluid flow between the inlet gallery and the cylinders of the pump. -
FIG. 8 graphically illustrates the relationship of the size of the open area, or flow area, of thecontrol apertures 95 versus the position of thethrottle plate 90, which is related to the linear position of theactuator piston 108 for the exemplary pump. The flow area of the orifice is directly related to the magnitude of the fluid flow there through. The actuator piston and the throttle plate move from a first position, corresponding with the orifice being fully (100 percent) open, to a second position, corresponding with the orifice being fully closed (0 percent open). A mid position is located halfway between the first and second positions, i.e., at 50 percent of the travel distance between the first and second positions. Broadly speaking, as the throttle plate moves from the first position to the second position, the average rate of change of the flow area of the variable orifice relative to movement of the actuator piston is significantly greater during the first half of travel compared to the second half of travel. For example, the flow area of the variable orifice created by the position of thecontrol aperture 95 relative to thetransmission aperture 94 decreases by at least 80 percent in the first 50 percent of the travel of the actuator piston from the initial first position, as depicted apoint 122 that occurs at the mid position inFIG. 3 . This rapid closure rate of the variable orifice occurs in what is referred to as afirst section 124 of the throttle plate rotation. - Thereafter the rate of change of the flow area decreases significantly slower, requiring that the
actuator piston 108 move the remaining 50 percent of the travel distance (thesecond section 126 of the throttle plate rotation) to reduce the flow area the final 20 percent to the fully closed position. Thus during the second section of throttle plate rotation, the piston and the throttle plate decreases the flow from 20 percent of the maximum flow to zero flow over the same amount (i.e., 50 percent) of throttle plate rotation as it takes to decrease the flow from 100 percent to 20 percent of the maximum flow. In other words, at a constant rate of rotation of thethrottle plate 90, the flow area of the variable orifice changes from the maximum flow area to about 20 percent of that maximum flow area at a rate that is at least twice as fast and the rate at which the flow area changes between about 20 percent of that maximum flow area and zero flow area. Therefore from the fully aligned aperture position, rotation of the throttle plate initially produces a relatively rapid decrease in flow area and then the flow area decrease occurs at a slower rate the as aperture motion approached the closed position. The inverse rates of change occur as thethrottle plate 90 moves clockwise in the drawings and the variable orifice, formed by the degree of alignment of thecontrol apertures 95 with thetransmission apertures 94, opens greater amounts. - The use of a
throttle plate 90 to control the amount of flow between theinlet gallery 31 and theinlet passages 26 enables the displacement of thepump 10 to be dynamically varied. When thethrottle plate apertures 95 are only partially aligned with the transitionplate transmission apertures 94, the amount of fluid flowing into thecylinder chamber 37 during the intake phase of each piston cycle is reduced. As a result, thepiston 52 reaches bottom dead center without thecylinder chamber 37 being completely filled with hydraulic fluid. Thus, a portion of the total effective piston displacement is lost. The amount of lost displacement does not vary significantly as a function of the speed of the pump, since the average pressure drop across the throttle plate is constant for typical pump speeds of 800 to 2500 RPM. - The present pump configuration with the
rotatable throttle plate 90 provides variable throttle choking at the input of each inlet check valve. This has a significant advantage over a pump that has throttle choking at a single place for all the cylinders, such as between theinlet port 28 and theinlet gallery 31. With the per inlet check valve choking arrangement of thepresent pump 10, the fluid volume between the throttle plate and the inlet check valve is relatively small and results in improved consistency and dynamic response in both starting and stopping fluid flow. - The foregoing description was primarily directed to a preferred embodiment of the invention. Although some attention was given to various alternatives within the scope of the invention, it is anticipated that one skilled in the art will likely realize additional alternatives that are now apparent from disclosure of embodiments of the invention. Accordingly, the scope of the invention should be determined from the following claims and not limited by the above disclosure.
Claims (27)
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US13/343,436 US8926298B2 (en) | 2012-01-04 | 2012-01-04 | Hydraulic piston pump with a variable displacement throttle mechanism |
CN201310002456.7A CN103195680B (en) | 2012-01-04 | 2013-01-04 | There is the hydraulic piston pump of variable displacement throttle mechanism |
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US13/343,436 US8926298B2 (en) | 2012-01-04 | 2012-01-04 | Hydraulic piston pump with a variable displacement throttle mechanism |
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Cited By (5)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US20140202325A1 (en) * | 2010-05-25 | 2014-07-24 | Husco International, Inc. | Compact Radial Piston Hydraulic Machine Having a Cylinder Block with Deforming Regions |
CN105089961A (en) * | 2014-05-06 | 2015-11-25 | 罗伯特·博世有限公司 | Hydrostatic piston machine |
EP3124783A1 (en) * | 2015-07-29 | 2017-02-01 | Hyundai Motor Europe Technical Center GmbH | High pressure pump |
US9746034B2 (en) | 2015-10-28 | 2017-08-29 | Deere & Company | Distributed load bearing with an inner flex ring |
CN109424593A (en) * | 2017-08-29 | 2019-03-05 | 佛山市科达液压机械有限公司 | A kind of constant pressure variable displacement pump transition plates |
Families Citing this family (6)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US9062665B2 (en) * | 2013-01-15 | 2015-06-23 | Husco International, Inc. | Hydraulic piston pump with throttle control |
GB2554005A (en) * | 2015-07-10 | 2018-03-21 | Husco Int Inc | Radial piston pump assemblies and use thereof in hydraulic circuits |
CN105952604A (en) * | 2016-06-28 | 2016-09-21 | 南京润泽流体控制设备有限公司 | Constant flow continuous injection pump |
CN109083821A (en) * | 2018-07-23 | 2018-12-25 | 江苏大学 | A kind of crankshaft connecting rod type wind energy suction function pump |
CN109630391B (en) * | 2018-12-17 | 2023-11-21 | 国家电网有限公司 | Adjustable throttling device capable of reducing flow disturbance |
CN112318842A (en) * | 2020-10-15 | 2021-02-05 | 合肥旭申信息科技有限公司 | Granulation equipment is retrieved to abandonment plastics cutlery box |
Citations (3)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US3434428A (en) * | 1967-06-05 | 1969-03-25 | White Motor Corp | Intake control for multiple piston pump |
US4065229A (en) * | 1976-10-01 | 1977-12-27 | General Motors Corporation | Variable capacity radial-4 compressor |
US6213729B1 (en) * | 1997-03-13 | 2001-04-10 | Luk Fahrzeung-Hydraulik Gmbh & Co., Kg | Suction-throttled pump |
Family Cites Families (11)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
CH385632A (en) | 1960-10-19 | 1964-12-15 | Schweizerische Lokomotiv | Axial piston pump with control device for changing the delivery rate |
US3418937A (en) | 1966-11-04 | 1968-12-31 | White Motor Corp | Radial piston pump |
WO1991002156A1 (en) * | 1989-08-05 | 1991-02-21 | Zahnradfabrik Friedrichshafen Ag | Piston pump |
US5634777A (en) | 1990-06-29 | 1997-06-03 | Albertin; Marc S. | Radial piston fluid machine and/or adjustable rotor |
CN1082143C (en) | 1993-11-08 | 2002-04-03 | Crt公共铁路技术公司 | Control device for a variable volume pump |
EP0690220A1 (en) | 1994-06-29 | 1996-01-03 | Lucas Industries Public Limited Company | Variable output pump |
JPH09228943A (en) | 1996-02-23 | 1997-09-02 | Nissan Motor Co Ltd | Oil hydraulic pump |
JPH11257239A (en) * | 1998-03-16 | 1999-09-21 | Unisia Jecs Corp | Variable displacement plunger pump |
JP2004176601A (en) | 2002-11-26 | 2004-06-24 | Komatsu Ltd | Capacity control device and positioning device for radial piston pump or motor |
DE10330757A1 (en) * | 2003-07-07 | 2005-02-03 | Bernhard-Rudolf Frey | Eccentric drive for volumetric pumps or motors |
CN101994673A (en) * | 2009-08-27 | 2011-03-30 | 潘建民 | Eccentric push circumferentially-distributed plunger pump |
-
2012
- 2012-01-04 US US13/343,436 patent/US8926298B2/en active Active
-
2013
- 2013-01-04 CN CN201310002456.7A patent/CN103195680B/en active Active
Patent Citations (3)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US3434428A (en) * | 1967-06-05 | 1969-03-25 | White Motor Corp | Intake control for multiple piston pump |
US4065229A (en) * | 1976-10-01 | 1977-12-27 | General Motors Corporation | Variable capacity radial-4 compressor |
US6213729B1 (en) * | 1997-03-13 | 2001-04-10 | Luk Fahrzeung-Hydraulik Gmbh & Co., Kg | Suction-throttled pump |
Cited By (6)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US20140202325A1 (en) * | 2010-05-25 | 2014-07-24 | Husco International, Inc. | Compact Radial Piston Hydraulic Machine Having a Cylinder Block with Deforming Regions |
CN105089961A (en) * | 2014-05-06 | 2015-11-25 | 罗伯特·博世有限公司 | Hydrostatic piston machine |
EP3124783A1 (en) * | 2015-07-29 | 2017-02-01 | Hyundai Motor Europe Technical Center GmbH | High pressure pump |
US9989027B2 (en) | 2015-07-29 | 2018-06-05 | Hyundai Motor Europe Technical Center Gmbh | High pressure pump having lubricating and cooling structure |
US9746034B2 (en) | 2015-10-28 | 2017-08-29 | Deere & Company | Distributed load bearing with an inner flex ring |
CN109424593A (en) * | 2017-08-29 | 2019-03-05 | 佛山市科达液压机械有限公司 | A kind of constant pressure variable displacement pump transition plates |
Also Published As
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US8926298B2 (en) | 2015-01-06 |
CN103195680B (en) | 2016-03-30 |
CN103195680A (en) | 2013-07-10 |
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